1. Field of the Invention
The present invention relates to a rotary compressor and, more specifically, to a compressor of the type having an electric motor and a compressing mechanism drivingly connected together by a crank shaft.
2. Description of the Prior Art
A conventional rotary compressor such as one shown in FIG. 13 has an electric motor 1 and a compressing mechanism 2 connected together by a crank shaft 10. A closed housing 14 accommodates these members.
The electric motor 1 is disposed in the closed housing 14 at an upper portion thereof and comprises a rotor 1a and a stator 1b. The crank shaft 10 has one end thereof fitted in the rotor 1a and drivingly connected at the other end to the compressing mechanism 2.
The compressing mechanism 2 comprises a cylinder block 3 fixed to an inner peripheral surface of the closed housing 14; a roller 4 constituting a rolling piston which is rotatably fitted on an eccentric portion 10a of the crank shaft 10, with the eccentric portion being positioned in a cylinder bore 3a formed in the cylinder block 3; a vane 5 capable of reciprocating as the roller 4 rotates about the axis of the crank shaft 10; main and auxiliary bearings 6 and 7 mounted on the cylinder block 3 in such a manner so as to close the upper and lower ends of the cylinder bore 3a, respectively, while supporting the crank shaft 10; a discharge valve 9 provided on the auxiliary bearing 7; and a cover 8 covering the discharge valve 9.
The compressor is equipped with balancing weights (hereinafter referred to as "balancers") for cancelling an offset force generated by the eccentric rotation of the rolling piston 4. These balancers comprise a first balancer 11-1 mounted on the end of the crank shaft 10 adjacent the auxiliary bearing 7, a second balancer 11-2 mounted on the lower end of the rotor 1a of the electric motor 1, and a third balancer 11-3 mounted on the upper end of the rotor 1a. The first balancer 11-1 is covered with a cover 12.
In FIG. 14 the amount M.sub.0 R.sub.0 of unbalance in the eccentric portion 10a of the crank shaft 10 is the total of values each obtained by multiplying the mass of each of the rotary portions of the eccentric portion 10a by its distance from the center of gravity.
M.sub.1 R.sub.1 is the total of values each obtained by multiplying the mass of each of the portions of the first balancer 11-1 by its distance from the center of gravity and is called an amount of balance. Similarly, M.sub.2 R.sub.2 is the amount of balance of the second balancer 11-2, and M.sub.3 R.sub.3 is the amount of balance of the third balancer 11-3.
As shown in FIG. 14, the respective axial distances from the position of the unbalance amount M.sub.0 R.sub.0 to the first, second and third balancers 11-1, 11-2 and 11-3 are represented by l1, l2 and l3, respectively.
Balance in forces is achieved in the manner expressed by the following equation: EQU M.sub.0 R.sub.0 +M.sub.3 R.sub.3 =M.sub.2 R.sub.2 +M.sub.1 R.sub.1 (1)
Balance in moments is achieved in the manner expressed by the following equation: EQU M.sub.1 R.sub.1 .multidot.l1+M.sub.3 R.sub.3 .multidot.l3=M.sub.2 R.sub.2 .multidot.l2 (2)
Balance in the primary vibration mode shown in FIG. 15 is achieved in the manner expressed by the following equation: EQU A.sub.1 .multidot.M.sub.1 R.sub.1 +A.sub.2 .multidot.M.sub.2 R.sub.2 =A.sub.0 .multidot.M.sub.0 R.sub.0 +A.sub.3 .multidot.M.sub.3 R.sub.3 (3)
where A.sub.0, A.sub.1, A.sub.2 and A.sub.3 express the coefficients of the primary vibration mode.
From the above three equations, the unknown balance amounts M.sub.1 R.sub.1, M.sub.2 R.sub.2 and M.sub.3 R.sub.3 can be calculated.
The above-described balancers in the crank shaft system are employed when the compressor operates at a frequency close to the primary critical speed of the shaft system. The balance amounts of these balancers are so determined as to cancell the primary vibration mode, i.e., the primary deflection mode.
The above-described structure is disclosed in Japanese Utility Model Unexamined Publication No. 59-107984.
The above-described prior art, however, gives no consideration to variations between individual balancer members. That is, when the tolerance level is determined in view of mass-productivity, variations may occur in the balance amounts of the balancers. Factors which can cause such variations include variations in the dimensions and density of the balancers per se, variation in the mounting angle caused during the mounting thereof, and variation in the amount of deformation caused during the assembly thereof.
Furthermore, variations may occur between rotors 1a of individual electric motors 1 in the level of eccentricity of their centers of gravity. Since the mass of the rotor 1a is relatively great, such variations can greatly influence shaft vibration.
FIG. 16 shows the deflection of the crank shaft determined in the case where the influence of the above-described various variation factors is taken into consideration. The one-dot-chain line 100 indicates the curve along which the axis of the crank shaft 10 deflects when the associated balancers and the associated rotor have no variations and are at the medians of their respective design values (the set values). The solid lines 101 and 102 indicate two different deflection curves of the axis of the crank shaft 10 obtained when there is an influence by variations. It will be understood from FIG. 16 that, when the amount of deflection is observed at the upper end of the rotor 1a, the deflection amount which results from variations is .delta.1 or .delta.2, whereas the amount of deflection is substantially negligible when the above-mentioned component parts are at the median values.
If deflection amounting to .delta.1 or .delta.2 occurs, the crank shaft 10 is inclined in the shaft hole of the main bearing 6 to cause a one-sided engagement therewith. This leads to a problem that such component parts as the main bearing 6 and the crank shaft 10 become worn to an abnormal extent. Another problem is that the deflection of the axis of the crank shaft 10 increases the level of vibration of the entire compressor to an abnormal extent.
FIG. 17 is a graph showing data useful to compare the vibration value, i.e., the vibration acceleration, of a crank shaft with balance amounts at the medians of the set values, and the vibration value of another crank shaft with balance amounts greatly varying from the set values. The vibration value of the first crank shaft is indicated by the solid line with circles therealong, while that of the second crank shaft is indicated by the broken line with squares therealong. As will be clearly understood from the graph, the crank shaft with great variations encounters a rapid increase in the vibration acceleration as its speed of rotation increases.